In certain piston-and-cylinder applications, it is frequently convenient to provide a tandem piston assembly in which two axially spaced effective surfaces, e.g. of different effective area, can be reciprocated on a common carrier in respective cylinders or working chambers.
This is the case, for example, in piston-type compressors using a stepped piston, i.e. a single body having a large diameter portion and a small diameter portion defining the respective surface areas and effective in respective chambers to compress a gas therein, e.g. for two-stage compression. In other words, the gas compressed by one of the pistons can be fed into the chamber of the other for increased compression.
Such stepped pistons are provided with piston rings and guide rings which mechanically bear upon the walls of the respective cylinders, serve to seal the clearance between the piston and the cylinder wall against leakage of gas, and of course, are subject to wear.
The problem is especially pronounced in so-called dry-running compressors, i.e. compressors in which the seal between each piston and the respective cylinder wall is not lubricated by a liquid. In this case, the pressure differential along the length of the piston is limited by the mechanical properties of the sealing ring and, of course, varies with wear of the sealing ring.
Furthermore, the pressure differential across the sealing ring has an effect on the mechanical properties and stability thereof since, by subjecting the sealing ring to elevated pressures, one applies stress which tends to increase the danger of rupture or bring about added wear.
As a consequence, to avoid rupture of the piston ring by overstressing of the latter, the pressure differential across the ring, especially in the high pressure stages of a dry-running compressor must be limited and this, for a given output pressure, means that the number of stages to achieve the desired pressure level must be increased.
Since an increased number of stages means increased wear, the maintenance activity associated with such compressors is considerable. Furthermore, as the number of stages increases, the energy losses increase and the construction becomes increasingly complex.
Both initial construction and operation, therefore, may be unduly complicated by the need for a large number of operative stages.
Mention may also be made of the fact that contactless pistons have been provided heretofore in the compressor art as well. In this case, the guiding of the piston rod utilizes a fixed guide bushing, sleeve or the like, disposed below the cylinder chamber. While here problems with piston rings are eliminated, this approach introduces an entirely different array of problems resulting from the manner in which the piston rod is guided and, indeed an insufficiency in the ability to properly guide the piston rod so as to avoid vibration or oscillation of the piston and the portion of the rod carrying the piston beyond the guide.
Since the piston rod is a spring with respect to transverse forces and the piston itself is a mass free to move laterally at the end of this rod, the entire assembly constitutes a vibratile element whose effective length varies as the piston rod moves with respect to the fixed guide, thereby changing the natural or resonance frequency and creating vibration problems which ultimately are detrimental to the guide and to the compressor as a whole. Accordingly, the use of sealless or contactless pistons under circumstances described has not successfully replaced the use of piston-ring systems in spite of the fact that these latter systems involve substantial problems.